DE69930181T2 - Internal combustion engine - Google Patents

Description

BACKGROUND OF THE INVENTION

1. FIELD OF THE INVENTION

The The invention relates to an internal combustion engine according to the preamble of claim 1, wherein for combustion an inert gas in a combustion chamber is initiated.

2. DESCRIPTION OF THE STAND OF THE TECHNIQUE

Around in an internal combustion engine such as a diesel engine to limit the production of NOx, the construction is conventional that a machine exhaust tract and an engine intake tract by an exhaust gas recirculation (below referred to as EGR) are connected to a waste gas and an EGR gas via the EGR tract to recirculate into the engine intake tract. There the EGR gas has a relatively high heat capacity and accordingly a big one heat In this case, the combustion temperature becomes within the combustion chamber with increasing amount of EGR gas or with increasing EGR ratio (RGR gas quantity / (EGR gas quantity + Intake air amount)). With lower combustion temperature decreases the amount of production of NOx. So the stronger the EGR ratio increases, the lower the generation amount of NOx.

It Thus, it is known that the generation amount of NOx is lowered at a higher EGR ratio can be. If the EGR ratio elevated but the production amount of soot or smoke starts to increase suddenly, if the EGR ratio exceeds a certain limit. As far as this point is concerned, it has been assumed that the smoke in a further increase in the EGR ratio unlimited increases, so that the EGR ratio, where the smoke suddenly begins to increase as the maximum allowable limit for the EGR ratio is considered has been.

Accordingly is the EGR ratio so far has been defined within a scope that does not have the maximum permissible Limit value goes out. The maximum permissible limit of the EGR ratio is dependent very different from the type of machine and the fuel, though it is within a range between about 30% and 50%. at a conventional one Diesel engine became the EGR ratio therefore limited to the range between about 30% and a maximum of 50%.

There it has now been assumed that it is with respect to EGR ratio the maximum allowable Limit value, the EGR ratio was not above the maximum permissible Limit outgoing range defined so that the production quantity the smoke as low as possible was. If the EGR ratio, however is defined so that the amount of production of NOx and smoke so as far as possible also has the reduction of the generation amount of NOx and smoke a limit. Therefore, a considerable amount remains generated at NOx and smoke.

Indeed is during the investigation of combustion in the diesel engine has been found that the smoke suddenly increases as described, if the EGR ratio over the maximum permissible In addition, limit value increases, but the amount of production of the smoke increases has a maximum and the smoke suddenly begins to decrease if the EGR ratio continues beyond the maximum value increased becomes. If the EGR ratio during the Idle to a value greater than or equal to set equal to 70 or the EGR gas is strongly cooled when the EGR ratio is on a value greater or greater is set equal to 55, the smoke is generally 0 or is hardly produces soot. Furthermore it has been established that at this time also the production quantity at NOx is much smaller. Based on this information then considerations to the cause why soot is not produced. As a result of which a new combustion system has been designed, so far not available yet is and at the same time soot and NOx can reduce. The new combustion system will be explained in detail below. Short It is said that the basic idea is the growth of hydrocarbon to stop before it grows to soot.

And although it has been confirmed by many experiments and investigations, that the growth of hydrocarbon ends before it becomes soot, if the temperature of the fuel and the ambient gas in the combustion chamber while the combustion is less than or equal to a certain temperature and that the hydrocarbon rapidly increases to soot if the temperature of the fuel and the ambient gas higher than the specific one addressed Temperature is. The temperature of the fuel and the ambient gas is in this case in the fuel combustion strong by a affects the endothermic effect of the gas surrounding the fuel, that's why it's possible is to control the temperature of the fuel and the ambient gas, by the heat absorption amount of the gas surrounding the fuel in accordance with the generation amount is set at the time of fuel combustion.

Accordingly, when the temperature of the fuel and the ambient gas during combustion within the combustion chamber is restricted to a level lower than or equal to the temperature at which the growth of the hydrocarbon ends up on the way, soot is not generated. It is therefore possible to restrict the temperature of the fuel and the ambient gas during combustion within the combustion chamber to a level less than or equal to the temperature at which the growth of the hydrocarbon ends by adjusting the heat absorption amount of the gas surrounding the fuel. On the other hand, the hydrogen whose growth ends before it becomes soot can be easily purified by a post-treatment using an oxidation catalyst or the like. This is the basic principle of the new combustion system. The internal combustion engine using the new combustion system is already under the number of the applicant EP 0 879 946 A been registered. Further, reference is made to Yanagihara H. et al .: "A study of diesel combustion under uniformly higher dispersed composition formation", JSAE Review 18 (1997), pages 247-254.

at the new combustion system is the fuel injection time, to which a stable combustion can be achieved, while a relatively tight Crank angle range limited. So if the fuel injection timing is preferred, the injected fuel through the compressed High temperature gas heated up long, so the temperature of the fuel and the ambient gas during the combustion gets high. As the hydrocarbon increases to soot, therefore smoke is generated. When the fuel injection timing against delayed becomes, the temperature of the injected fuel is not so very elevated, so most of the fuel does not burn, resulting in a misfire. at This new combustion system therefore has an optimum range of crank angles. in which it related to the fuel injection timing neither Smoke still to the misfire comes, why it for the case that the new combustion is performed is necessary, the Carry out fuel injection in the optimum crank angle range.

However, it changes the optimum crank angle range in which a stable combustion can be achieved easily with the environmental conditions, such as with the climate. By the change The environmental conditions are therefore also smoke or too Misfires when the fuel injection timing is within the optimum crank angle range is controlled in advance according to the operating state of the machine is determined.

Apart from that, on the DE 196 12 179 C1 which discloses a method of controlling combustion in a conventional internal combustion engine. For this purpose, the combustion noise of individual cylinders is measured by means of a structure-borne sound sensor and the measured value of each cylinder is compared with a threshold value. If the threshold is exceeded, it is decided that combustion has started in the respective cylinder. By advancing or retarding the injection timing, a different torque output of the individual cylinders is compensated, which is caused by machining tolerances in the manufacture of the internal combustion engine.

SUMMARY THE INVENTION

task The invention is to carry out a new type of combustion, the neither smoke nor misfires generated.

The The above object is achieved by the feature combination of the independent claim solved. The dependent claims disclose further advantageous embodiments of the invention.

The The above object is achieved by the invention achieved that for an internal combustion engine is provided, which is constructed so that a generation amount of soot gradually a maximum increases as the amount of an inert gas supplied within a combustion chamber increases, being provided with an inert gas amount control means for limitation the temperature of a fuel and an ambient gas at the time combustion within the combustion chamber to a temperature which is lower than a temperature when soot is generated by the amount of the inert gas supplied within the combustion chamber to more than the amount of inert gas is set when the production amount of soot maximum value a detecting means for detecting a torque change amount the machine power and a control device for control of the fuel injection timing, so that the amount of torque change within a predetermined range.

Of Further, in the above internal combustion engine, the injection control device According to the invention, the fuel injection time considering the detected torque change amount Taxes.

Moreover, in the above internal combustion engine, according to the present invention, the injection timing control means may change the fuel injection timing depending on the detected torque change amount so that the injection timing is advanced or retarded when the detected torque change amount not within the predetermined range.

In addition, can in the above internal combustion engine, the injection timing control means According to the invention, the fuel injection time if the amount of torque change greater than becomes the predetermined range, and the fuel injection timing delay, when the torque change amount becomes smaller than the predetermined range.

Of Further, in the above internal combustion engine, the detection means According to the invention a mean amount of torque change of all cylinders and the injection timing controller control the fuel injection timing of each cylinder so that the mean amount of torque change within the predetermined range. Instead, you can the detector also detects the amount of torque change on each cylinder and the injection timing control means detects the fuel injection timing on each cylinder so control that the amount of torque change each cylinder is within the predetermined range.

If As the combustion deteriorates, the torque change increases the machine performance, which is why it is possible, based on the Torque change amount To judge the machine performance, whether the combustion deteriorates or not. When the fuel injection timing is advanced will take, as well Torque variation amount the engine power gradually however, smoke is generated when the fuel injection timing is too much preferred. When the torque change amount of engine power Controlled to a specific area, so can be an improved Achieve combustion that produces neither misfire nor smoke.

These Brief description of the invention does not necessarily describe all necessary Features, why the invention in the form of a sub-combination these described features may be present.

SUMMARY THE DRAWINGS

1 shows a general view of a compression ignition type of internal combustion engine according to the invention;

2 graphically shows the change of the throttle opening degree, the EGR ratio, the output torque and the respective amounts of smoke, HC, CO and NOx according to the air-fuel ratio;

3A and 3B represent as a graph the combustion pressure as a function of the air-fuel ratio;

4 shows the molecular structure of the fuel;

5 graphically shows the relationship between the generation amount of soot and the EGR ratio;

6 graphically shows the relationship between the total amount of intake gas and the required load;

7 shows graphically a first operating range I and a second operating range II;

8th graphically shows the output of an air-fuel ratio sensor;

9 graphically shows an opening degree of a throttle valve, an opening degree of an EGR control valve, an EGR ratio, an air-fuel ratio, an injection timing, and an injection amount at the demanded load;

10 graphically shows an air-fuel ratio in a first operating region I;

11A and 11B show a map of a target opening degree of the throttle valve and the like;

12A and 12B show a map of an injection amount Q and an injection start timing θS;

13 graphically shows an air-fuel ratio in a second combustion mode;

14A and 14B show a map of a target opening degree of the throttle valve and the like;

15A and 15B show a map of the injection amount Q and the injection start timing θS;

16 shows the change of the crank angle over time;

17 shows a flowchart for calculating the deviation of the torque in the first embodiment;

18 shows a flowchart for controlling the engine operation in the first embodiment example;

19 shows a flowchart for calculating the deviation of the torque in the second embodiment;

20 shows a flowchart for controlling the engine operation in the second embodiment;

21A and 21B show a relationship between a pressure on an accelerator pedal, a rotational speed, and an injection amount for decreasing the change of the output torque when switching the combustion state;

22A and 22B show a map of an injection amount for reducing the change of the output torque when switching the combustion state;

23A and 23B show the change in the injection amount for reducing the change of the output torque when switching the combustion state;

24 shows a map of the injection quantity;

25 show the first operating region I and the second operating region II for switching the combustion on the basis of the required torque;

26A and 26B show as a graph the required torque;

27 shows an air-fuel ratio in the first operating region I for switching the combustion based on the required torque;

28A and 28B show a map of the injection amount and the injection start timing for switching the combustion based on the requested torque;

29A and 29B and 29C show a map of the target opening degree of the throttle valve and the EGR control valve and the target fuel pressure for switching the combustion based on the required torque;

30 shows an air-fuel ratio in the second operating region II for switching the combustion based on the required torque;

31A and 31B show a map of the injection amount and the injection start timing for switching the combustion based on the requested torque;

32A and 32B and 32C show a map of the target opening degree of the throttle valve and the EGR control valve and the target fuel pressure for switching the combustion based on the required torque;

33 shows a flowchart for controlling the engine operation to switch the combustion based on the required torque; and

34 FIG. 10 is a flowchart for controlling the engine operation to switch the combustion based on the requested torque. FIG.

DESCRIPTION THE PREFERRED EMBODIMENTS

1 Fig. 14 shows a case where the invention is applied to a compression-ignition type four-cycle internal combustion engine.

In 1 denotes the reference number 1 a machine body, the reference number 2 a cylinder block, the reference number 3 a cylinder head, the reference number 4 a piston, the reference number 5 a combustion chamber, the reference number 6 an electrically controlled fuel injection valve, the reference numeral 7 an inlet valve, the reference number 8th an inlet channel, the reference number 9 an exhaust valve and the reference number 10 an outlet channel. The inlet channel 8th is over an end inlet branch pipe 11 with a surge tank 12 connected while the expansion tank 12 via an inlet pipe 13 and an intercooler 14 with a supercharger, for example an outlet section of a compressor 16 an exhaust gas turbocharger 15 , connected is. The inlet section of the compressor 16 is above an air intake pipe 17 with an air filter 18 in conjunction, being inside the air intake pipe 17 one from a stepper motor 19 driven throttle valve 20 is arranged.

The outlet channel 10 On the other hand, there is an exhaust manifold 21 and an exhaust pipe 22 with the inlet section of an exhaust gas turbine 23 the exhaust gas turbocharger 15 in connection, while the exhaust section of the exhaust gas turbine 23 over an exhaust pipe 24 with a cat 26 communicating in itself a catalyst 25 containing oxidation function. Inside the exhaust manifold 21 is an air-fuel ratio sensor 27 arranged.

One with the outlet section of the cat 26 connected exhaust pipe 28 and the downstream of the throttle valve 20 located air intake pipe 17 are together via an AGR tract 29 being within the EGR tract 29 one from a stepper motor 30 driven EGR control valve 31 is arranged. Furthermore, within the AGR tract 29 an intercooler 32 for cooling one in the AGR tract 29 arranged flowing EGR gas. At the in 1 shown embodiment is in the intercoolers 32 an engine cooling water is introduced and the EGR gas is cooled by the engine cooling water.

Furthermore, the fuel injection valve 6 via a fuel supply pipe 33 with a fuel storage, a so-called common rail 34 connected. The fuel becomes the common rail 34 via an electrically controlled fuel pump 35 supplied, the discharge amount is changeable. The the common rail 34 supplied fuel is the fuel injector 6 over the respective Kraftstoffzuführrohr 33 fed. In the common rail 34 is a fuel pressure sensor 36 for detecting the fuel pressure within the common rail 34 built-in, so that the discharge amount of the fuel pump 35 can be controlled so that the fuel pressure within the common rail 34 based on an output signal of the fuel pressure sensor 36 becomes a target fuel pressure.

As in 1 is shown is also a crankshaft 37 with an automatic transmission 38 connected.

An electronic control unit 40 consists of a digital computer and is equipped with a read-only memory (ROM) 42 , a random access memory (RAM) 43 , a microprocessor (CPU) 44 , an input block 45 and an output module 46 provided to each other via a bidirectional bus 41 are connected. The output signal from the air-fuel ratio sensor 27 and the output signal from the fuel pressure sensor 36 are each via a corresponding AD converter 47 into the input block 45 entered. With an accelerator 50 is a load sensor 51 for generating an output voltage proportional to the amount of depression L of the accelerator pedal 50 connected, the output voltage of the load sensor 51 via the corresponding AD converter 47 into the input block 45 is entered. Furthermore, with the input block 45 a crank angle sensor 52 connected, each time when the crankshaft rotates by 30 degrees, for example, generates a dispensing pulse. In addition, in the input block 45 a vehicle speed indicative output pulse from the vehicle speed sensor 53 and an output of a neutral sensor 54 entered, which detects whether the automatic transmission 38 in a neutral position. It also enters the input block 45 an output signal of an idle switch 55 entered, which detects whether the position of the throttle valve 20 corresponds to an idling opening degree or not. In contrast, the output device is 46 via the output module 46 corresponding drive circuit with the fuel injection valve 6 , the stepper motor 19 for controlling the throttle valve, the stepping motor 30 for controlling the EGR control valve and the fuel pump 35 connected.

2 FIG. 12 shows an experimental example indicating the change of the output torque and the change of the discharge amount of smoke, HC, CO and NOx when the air-fuel ratio A / F (abscissa axis in FIG 2 ) is changed by the opening degree of the throttle valve 20 and the EGR ratio are adjusted. As it turned out 2 If the air-fuel ratio is less than or equal to the stoichiometric ratio, the smaller the air-fuel ratio A / F is, the larger the EGR ratio becomes in this experimental example, and the EGR ratio is greater than or equal to 65% Air-fuel ratio is 14.6.

As in 2 In the case of reducing the air-fuel ratio A / F, by increasing the EGR ratio, the generation amount of smoke starts to increase as the EGR ratio approaches 40% and the air-fuel ratio A / F approaches approximately 30% reached. Then, when the EGR ratio is further increased and the air-fuel ratio A / F is lowered, the generation amount of smoke suddenly increases until it reaches a maximum value. Then, when the EGR ratio is further increased and the air-fuel ratio A / F is lowered, the generation amount of smoke abruptly drops and the generation amount of smoke reaches when the air-fuel ratio A / F is adjusted while the air-fuel ratio A / F is set EGR ratio approaches greater than or equal to 75% 15.0, close to 0. Thus, soot is hardly produced. At this time, the output torque of the engine is a little, and the generation amount of NOx is much lower. The respective production amount of HC and CO start to rise at this time against it.

3A shows the change in the combustion pressure within the combustion chamber 5 when the air-fuel ratio A / F is close to 21 and the generation amount of smoke is greatest while 3B the change in combustion pressure within the combustion chamber 5 When the air-fuel ratio A / F is close to 18, the generation amount of smoke is large and grain 0 is 0. As can be seen from the comparison between the 3A and 3B results, the combustion pressure in the case of 3B in which the amount of smoke produced by and large is 0 is smaller than in the case of 3A in which the generation amount of smoke is large.

From the into the 2 and 3 The following conclusions can be drawn from the test results shown. Namely, first, the generation amount of NOx, as in 2 is significantly lower when the air-fuel ratio A / F is equal to or smaller than 15.0, and the generation amount of smoke is generally 0. The smaller generation amount of NOx means a reduction in the combustion temperature within the combustion chamber 5 , which is why it can be said that the combustion temperature within the combustion chamber 5 low is when hardly soot is produced. The same can be said for 3 say. And that is in the state of 3B , in which hardly soot is produced, the combustion pressure is low, so that the combustion temperature within the combustion chamber 5 is low.

Second, as in 2 2, the discharge amount of HC and CO is shown when the generation amount of smoke and the generation amount of soot are, by and large, zero. This means that the hydrocarbon is released without growing into soot. Namely, in a straight-chain hydrocarbon or an aromatic hydrocarbon, as in 4 is shown and contained in the fuel, due to thermal decomposition, a precursor of the soot is formed when the temperature rises in a state of oxygen deficiency, and soot resulting from a solid mainly composed of an aggregation of carbon atoms is formed. The actual production process of the soot is complex in this case and it remains undetermined which state the precursor of the soot occupies, but the in 4 In any case, shown hydrocarbon via the precursor of the carbon black to the soot. Therefore, as mentioned above, when the generation amount of soot is generally 0, the discharge amount of HC and CO as in FIG 2 shown, wherein the HC at this time corresponds to the precursor of the carbon black or the hydrocarbon in the state before.

If the considerations based on in the 2 and 3 As a whole, the amount of production of soot is 0 when the combustion temperature is within the combustion chamber 5 is low, leaving the combustion chamber 5 the precursor of the carbon black or the hydrocarbon are discharged in the state before. Further detailed experiments and investigations have shown that the growth of the soot ends on the way or no soot is produced, if the temperature of the fuel and the ambient gas within the combustion chamber 5 are less than or equal to a certain temperature, and that soot is generated when the temperature of the fuel and the ambient gas within the combustion chamber 5 are greater than or equal to a certain temperature.

There the temperature of the fuel and the ambient gas to the Time when the growth of the hydrocarbon with the state of Precursor of the soot ends or the above-mentioned specific temperature due various causes such as the type of fuel, the compression ratio of Air-fuel ratio and like that changes, let yourself in this case do not say how high the temperature is. Indeed the particular temperature is closely related to the amount of production at NOx, so that the certain temperature to a certain extent can be determined based on the amount of production of NOx. In fact decreases the temperature of the fuel and the ambient gas at the time of combustion with increasing EGR ratio, so that the generation amount of NOx decreases. It hardly gets Generates soot, when the generation amount of NOx is close to 10 ppm or less. Accordingly falls the above specific temperature on the whole with the temperature together in which the generation amount of NOx is almost 10 ppm (g / t) or less.

Once soot has been produced, the soot can not be cleaned by the post-treatment which utilizes the oxidation function catalyst. On the other hand, by the post-treatment using the catalyst having an oxidation function, the precursor of the carbon black or the hydrocarbon of the state before can be easily purified. In view of the post-treatment by the oxidation-function catalyst, therefore, there is a great difference between the discharge of the hydrocarbon from the combustion chamber 5 as precursor of the soot or in the state before and the discharge of the hydrocarbon from the combustion chamber 5 as soot. The new combustion system used in this invention is primarily designed to remove the hydrocarbon from the combustion chamber 5 as a precursor of the soot or in the state before giving off without being inside the combustion chamber 5 Soot is generated, and the hydrocarbon is oxidized by the catalyst having the oxidation function.

In order for the growth of the hydrocarbon to finish in the state before the production of the soot, the temperature of the fuel and the ambient gas at the time of combustion must be within the combustion chamber 5 be limited to a temperature lower than the temperature at which the soot is produced. It is clearly understood in this case that the endothermic effect of the gas, which occurs when the fuel is burnt is located around the fuel, which affects the limitation of the temperature of the fuel and the ambient gas to a considerable extent.

If namely around the fuel only air is present, the vaporized reacts Fuel immediately with the oxygen in the air and burn. In this case increased the temperature of the air away from the fuel is not so much but it comes only to a significant local temperature rise around the fuel. That means that at this time the air off the fuel hardly any endothermic effect on the combustion heat in the fuel exerts. Since it is local in this case to a significantly increased Combustion temperature comes, produces unburned hydrocarbon, on the heat of combustion works, the soot.

If in contrast, the fuel is present in a gas mixture, the out of a large crowd Inert gas and a small amount of air are the conditions something else. In this case, the burnt fuel diffuses outward and reacts under combustion with the one mixed in the inert gas Oxygen. Because the heat of combustion in this case in the outer inert gas absorbed, increased the combustion temperature is not so much. The combustion temperature let yourself so restrict it to a low level. The presence of the Inert gas therefore plays an important role in the combustion temperature restrict, where possible is the combustion temperature due to the endothermic effect of the inert gas to a low level.

In order to limit the temperature of the fuel and the ambient gas to a lower temperature than the temperature at which soot is generated, an amount of inert gas sufficient to absorb enough heat of combustion is necessary in this case. As the amount of fuel increases, so does the necessary amount of inert gas. In this case, in this case, the larger the heat capacity of the inert gas, the stronger the endothermic effect, so that a gas having a large heat capacity is preferable for the inert gas. In view of this, since CO ₂ and the EGR gas have a relatively large heat capacity, it is preferable to use the EGR gas as the inert gas.

5 FIG. 14 shows the relationship between the EGR ratio and the smoke when the EGR gas is used as the inert gas and the degree of cooling of the EGR gas is changed. Namely, the curve A in 5 In the case of strongly cooling the EGR gas to keep the EGR gas temperature largely at 90 ° C, the curve B shows the case that the EGR gas is cooled by a compact refrigerator, and the curve C the case a lack of forced cooling of the EGR gas.

Like the curve A in 5 For example, when the EGR ratio is slightly less than 50%, the generation amount of soot becomes the highest in the case of strongly cooling the EGR gas, in which case hardly soot is generated when the EGR ratio is at a level of On the whole, greater or equal to 55% is set. As opposed to curve B in FIG 5 In the case of low EGR gas cooling, the generation amount of soot peaks when the EGR ratio is slightly larger than 50%, in which case hardly any soot is produced when the EGR ratio is at a level of By and large, greater or equal to 65% is set.

As further the curve C in 5 In the case of lack of forced cooling of the EGR gas, the generation amount of soot peaks when the EGR ratio is close to 55%, and hardly produces soot in this case when the EGR ratio is at a level of .mu.m Big and whole is set to greater than or equal to 70.

5 in this case, the generation amount of soot when the engine load is relatively high, with the EGR ratio, in which the production amount of soot peaks, and the lower limit of the EGR ratio, in which hardly soot is generated, slightly reduce when the engine load decreases. The lower limit of the EGR ratio at which hardly soot is produced changes as mentioned above according to the degree of cooling of the EGR gas and the engine load.

6 in the case where the EGR gas is used as the inert gas, the mixed gas amount of the EGR gas and the air necessary to make the temperature of the fuel and the ambient gas lower than the temperature at the time of combustion, in which the soot is generated, the proportion of air in the gas mixture and the proportion of the EGR gas in the gas mixture. The ordinate axis is in 6 for the total amount of intake gas entering the combustion chamber 5 is drawn in, wherein the dashed line Y indicates the total amount of intake gas, the combustion chamber 5 can pick up when no charge occurs. Furthermore, the abscissa axis shows the required load.

In 6 indicates the proportion of air or the amount of air in the gas mixture, the amount of air needed for a complete combustion of the injected fuel. The ratio between the amount of air and the fuel injection amount thus corresponds in the case of 6 the stoichiometry air-fuel ratio. In 6 is the proportion of the EGR gas or the amount of EGR gas in the gas mixture, the EGR gas amount that is required at least to adjust the temperature of the fuel of the ambient gas to a lower temperature than the temperature at which forms soot when injected fuel is burned. The EGR gas quantity is greater than or equal to 55% of the EGR ratio and that of the in 6 shown embodiment is greater than or equal to 70%. So if the whole in the combustion chamber 5 let inlet gas quantity on the solid line X in 6 and the ratio between the amount of air and the EGR gas amount within the total intake gas amount X to the in 6 are set, the temperature of the fuel and the ambient gas are lower than the temperature at which the soot is generated, and accordingly, no complete soot is generated. Further, the generation amount of soot at this time is about 10 ppm (g / t) or less, and therefore the generation amount of NOx is significantly lower.

Since the heat of combustion in the combustion of fuel increases as the amount of fuel injection increases, the amount of absorption of the heat due to the EGR gas must be increased to keep the temperature of the fuel and the ambient gas lower than the temperature at which the soot is generated. Accordingly, as in 6 is shown, the EGR gas amount are increased according to the fuel injection amount. The EGR gas quantity should therefore be increased if the required load is high.

If no charge, the upper limit for the amount X of the total is in the combustion chamber 5 let inlet gas Y, so that in the area in 6 in that the required load is greater than L ₀ , the air-fuel ratio can not be maintained at the stoichiometric air-fuel ratio as long as the EGR gas ratio is not reduced in accordance with the increasing required load. Thus, if it is intended to maintain the air-fuel ratio in the desired load range of more than L ₀ without stoichiometric air-fuel ratio charging, the EGR ratio is lowered in accordance with the higher required load, and therefore, in FIG it is impossible for the region with the desired load of more than L ₀ to maintain the temperature of the fuel and ambient gas at the temperature lower than the temperature at which the soot is produced.

However, if the EGR gas is in the range with the required load of more than L ₀ , as in 1 is shown, to the inlet side of the supercharger or via the EGR tract 29 in the air inlet pipe 17 the exhaust gas turbocharger 15 the EGR ratio can be maintained at a level of greater than or equal to 55%, for example 70%, which makes it possible to maintain the temperature of the fuel and the ambient gas at a temperature lower than the temperature, in which the soot is produced. So if the EGR gas is recycled so that the EGR ratio within the air intake pipe 17 70, for example, is also the EGR ratio of the intake gas as long as the compressor 16 can increase the pressure with which through the compressor 16 the exhaust gas turbocharger 15 increased pressure 70%, so that the temperature of the fuel and the ambient gas can be maintained at a temperature at which the soot is not generated. It is therefore possible to expand the operating range of the engine that can produce the low-temperature combustion.

When the EGR ratio is set to a level of greater than or equal to 55% in the range with the required load of more than L ₀ , the EGR control valve is 31 fully open in this case or the throttle valve 20 slightly closed.

As mentioned above, shows 6 the case of the combustion of the fuel under the stoichiometric air-fuel ratio, but the generation amount of NOx can then be limited to about 10 ppm or less, while the production of soot is limited when the amount of air to a lower level than the in 6 shown air quantity or the air-fuel ratio is set to rich or even if the amount of air to a higher level than the in 6 set air quantity or a lean average value of the air-fuel ratio, about 17 to 18 is set.

If the air-fuel ratio is made in a fuel surplus, but the excess fuel does not grow to the soot, because the combustion temperature is limited to a low temperature, so that no soot is generated becomes. Furthermore At this time, only a significantly smaller amount of NOx will be released generated. If the average air-fuel ratio is lean or if that Air-fuel ratio stoichiometric Air-fuel ratio on the other hand, a small amount of soot is produced if the combustion temperature is high. However, since the combustion temperature according to the invention to a low temperature limited will, will ever no soot produced. Furthermore Only a significantly smaller amount of NOx is produced.

When the low-temperature combustion is performed, so regardless of the air-fuel ratio, that is, regardless of whether If the air-fuel ratio is rich or stoichiometric, or if the average air-fuel ratio is lean, no soot is generated, so that the generation amount of NOx is significantly lower. Accordingly, in view of the improvement in the specific fuel consumption, it can be said that it is preferable to make the average air-fuel ratio lean.

This Case is on a machine operation with medium or low Load limited, at which the heat of combustion due to the combustion relatively low is about the temperature of the fuel and the ambient gas to Time of combustion within the combustion chamber to a level of less than or equal to the temperature limit at which the growth of the hydrocarbon ends up on the way. In the embodiment of the invention Therefore, the temperature of the fuel and the ambient gas at the time of combustion in a medium-duty engine operation or lower load to a temperature of less than or equal to the temperature is limited, at the growth of the hydrocarbon ends up on the way, so the perform first combustion mode or the low-temperature combustion, wherein in a high load engine operation, the second combustion mode or the conventional Combustion performed becomes. As from the above explanation indicates the first combustion mode or the low-temperature combustion in this case for a combustion in which the amount of inert gas within the Combustion chamber larger than the amount of the inert gas is at which the generation amount of the soot the maximum value occupies, and hardly produces soot becomes. The second combustion mode or the conventional combustion stands for combustion, in which the amount of inert gas within the combustion chamber is smaller is the amount of the inert gas at which the generation amount of the soot the highest value occupies.

7 FIG. 15 shows a first operation region I in which the first combustion mode is performed, and a second operation region II in which the second combustion mode is performed according to the conventional combustion method. The ordinate axis L in 7 in this case, the amount of depression of the accelerator pedal 50 or the required load and the abscissa axis N the speed. Furthermore, X (N) is in 7 a first boundary between the first operation area I and the second operation area II and Y (N) indicates a second boundary between the first operation area I and the second operation area II. The change of the operating range from the first operating range I and the second operating range II is decided on the basis of the first limit X (N) and the change of the operating range from the second operating range II to the first operating range I on the basis of the second limit Y (N).

If So the required load L the a function of speed N corresponding limit exceeds X (N) and when the operating condition of the machine is in the first operating range I and the low temperature combustion is performed, it is decided that the operating area to the second operating range II moves, so the combustion according to the conventional Combustion process performed becomes. If the required load L is now a function of Speed N corresponding second limit Y (N) falls below, is decided that the operating area to the first operating range I shifts, so that the low-temperature combustion is performed again.

The both above limits, the first limit X (N) and the closer to low load lying second boundary Y (N), it gives off the two following reasons. And that is one reason that, because of that, because the combustion temperature in the second operating range II on the high load side relatively high is, the low-temperature combustion can not be done immediately, even if the required load L at this time under the first boundary X (N) falls. That's because the low temperature combustion only then begins when the required load L significantly lower or lower than the second boundary Y (N). The second reason is, that in terms of change the operating range between the first operating range I and the second operating area II for a hysteresis is taken care of.

In this case, when the engine operating state is in the first operating region I and the low-temperature combustion is being performed, soot is hardly generated and instead is released from the combustion chamber 5 unburned hydrocarbon as a precursor of the carbon black or ejected in the state before. At the same time, that of the combustion chamber 5 discharged unburned hydrocarbon well through the catalyst 25 oxidized with the oxidation function.

As the catalyst 25 For example, an oxidation catalyst, a three-way catalyst or a NOx absorber can be used. The NOx absorber has the function of absorbing NOx when the average air-fuel ratio within the combustion chamber 5 is lean, and release NOx when the average air-fuel ratio within the combustion chamber 5 gets fat.

The NOx absorber is constructed, for example, such that alumina is used as a carrier and at least one element selected from an alkali metal such as potassium K, sodium Na, lithium Li and cesium Cs, an alkaline earth metal such as barium Ba and calcium Ca and a rare earth metal such as lanthanum La and yttrium Y, and a noble metal such as Platinum Pt is worn.

Like the oxidation catalyst, the three-way catalyst and the NOx absorber also have an oxidation function, and therefore, as mentioned above, the three-way catalyst and the NOx absorber as the catalyst 25 can be used.

8th shows the output of the air-fuel ratio sensor 27 , As in 8th is shown, the output current I of the air-fuel ratio sensor changes 27 according to the air-fuel ratio A / F. Thus, the air-fuel ratio based on the output current I of the air-fuel ratio sensor 27 be learned.

Next, referring to 9 summarized the operation control in the first operating range I and described in the second operating range II.

9 shows based on the required load L the degree of opening of the throttle valve 20 , the opening degree of the EGR control valve 31 , the EGR ratio, the air-fuel ratio, the injection timing and the injection amount. As in 9 is shown, in the first operating region I with the low required load L, the opening degree of the throttle valve 20 from an almost fully closed state gradually increases with increase in the required load L up to an opening degree of about two-thirds, while the opening degree of the EGR control valve 31 is increased from an almost fully closed state gradually with increase of the required load L to a fully opened state. In the in 9 In addition, in the embodiment shown, the EGR ratio is set to approximately 70% in the first operating region I and the air-fuel ratio is set to approximately the stoichiometric air-fuel ratio.

In other words, the opening degree of the throttle valve 20 and the opening degree of the EGR control valve 31 controlled in the first operating range I so that the EGR ratio is by and large 70% and the air-fuel ratio becomes a somewhat lean air-fuel ratio. At the same time, in this case, the air-fuel ratio by a correction of the opening degree of the EGR control valve 31 based on the output of the air-fuel ratio sensor 27 controlled to a lean target air-fuel ratio. Furthermore, the fuel injection takes place in the first operating region I before the upper compression dead center OT. In this case, the injection start timing θS is delayed in accordance with the increasing demanded load L, and the injection end timing θE is also delayed in accordance with the delayed injection start timing θS.

Further, at the time of idling, the throttle valve becomes 20 and at the same time also the EGR control valve 31 almost completely closed. When the throttle valve 20 is almost closed, the pressure inside the combustion chamber 5 low at the beginning of compression, so that the compression pressure is low. When the compression pressure is low, the compression work is by the piston 4 low, so that the vibrations of the machine main body 1 reduce. At the time of idling, the throttle valve becomes 20 so almost closed to the vibrations of the machine main body 1 to restrict.

However, when the operating range of the engine changes from the first operating region I to the second operating region II, the opening degree of the throttle valve becomes 20 suddenly increased from about the opening degree two-thirds toward fully open. At this time, the EGR ratio in the in 9 In the embodiment shown, the overall reduction is suddenly from 70% to 40% or less and the air-fuel ratio is suddenly increased. Since the EGR ratio is therefore over the range of the EGR ratio ( 5 ), in which a lot of smoke is generated, not so much smoke is generated when the operating range of the engine changes from the first operating region I to the second operating region II.

In the second operating region II, the second combustion mode or the conventional combustion is performed. According to this combustion method, a small amount of soot and NOx are generated, but the heat efficiency is higher than that of the low-temperature combustion, so that the injection quantity as shown in FIG 9 is suddenly reduced as the operating range of the engine changes from the first operating range I and second operating range II. In this second operating region II, the throttle valve 20 held to a portion thereof in the fully opened state and the opening degree of the EGR control valve 31 gradually decreases as the required load L increases. In addition, in this operating range II, the EGR ratio becomes low as the demanded load L increases and the air-fuel ratio becomes small as the required load L increases. However, the air-fuel ratio is set to a lean air-fuel ratio even if the gefor increased load L increases. Further, the injection start timing θS in the second operating region II is set near the compression top dead center OT.

10 shows the air-fuel ratio A / F in the first operating region I. In 10 The curves labeled A / F = 15.5, A / F = 16, A / F = 17 and A / F = 18 respectively indicate the states with air-fuel ratios of 15.5, 16, 17 and 18, respectively wherein the air-fuel ratios between the curves are each defined by proportional allocation. As in 10 is shown, the air-fuel ratio in the first operating region I is lean and the air-fuel ratio A / F is further made lean in the first operating region I in accordance with the decreasing required load L.

With decreasing required load L so the heat of combustion is lowered. Accordingly, the low-temperature combustion can be performed even when the EGR ratio is lowered as the required load L decreases. With decreasing EGR ratio, the air-fuel ratio increases, so that the air-fuel ratio A / F as in 10 shown with decreasing required load L becomes larger. With increasing air-fuel ratio A / F, the specific fuel consumption improves, so that the air-fuel ratio A / F according to the embodiment of the invention with increasing demanded load L is greater to the air-fuel ratio as lean as to make possible.

In this case, the target opening degree ST of the throttle valve becomes 20 needed to set the air-fuel ratio to the in 10 set target air-fuel ratio, previously in the ROM 42 in the form of a like in 11A map shown as a function of the required load L and the rotational speed N, while the target opening degree SE of the EGR control valve 31 needed to set the air-fuel ratio to the in 10 set target air-fuel ratio, previously in the ROM 42 in the form of a like in 11B map shown as a function of the required load L and the rotational speed N is stored.

In addition, the injection amount Q is controlled in performing the first combustion mode according to the required load L and the rotational speed N, wherein the injection amount Q previously within the first ROM 42 in the form of a like in 12A map shown as a function of the required load L and the rotational speed N is stored. Moreover, by trial, the optimum injection start timing θS is pre-calculated when the first combustion mode is performed, and the optimum injection start timing θS is previously in the ROM 42 in the form of a like in 12B map shown as a function of the required load L and the rotational speed N stored.

There the machine however during idle operation of the engine tends to oscillate the throttle valve and also the EGR control valve at the time of Idle mode almost closed. So according to the required Load an optimal amount of the EGR gas can be achieved have to For example, the speed and the negative pressure of the intake air pipe be controlled to a specific target value to the machine vibrations while of idle operation.

In order to limit the vibrations, it is then necessary to set the speed to a specific target speed, the negative pressure of the intake pipe at a certain target negative pressure, the air-fuel ratio to a specific target air-fuel ratio and the pressure change inside the combustion chamber 5 to set to a specific target pressure change. However, there is a close relationship between these parameters, therefore, in a controller in which one parameter is first set to a target value and next another parameter is set to a target value, the actual value of the already controlled parameter changes, and hence it becomes necessary is to control it again so that the parameter is set to the target value. Thus, to set the actual value of each parameter to the target value, each of the parameters must be controlled according to a specific set order.

So now when the idling operation, it is preferable, first the opening degree ST of the throttle valve 20 to correct so that the actual speed is set to the target speed, and next the opening degree SE of the EGR valve 31 to correct the negative pressure of the intake pipe to the target negative pressure of the intake pipe. However, when the negative pressure of the intake pipe is set to the target negative pressure of the intake pipe, the actual amount of intake air changes, so that the rotational speed changes. However, this change is smaller than, for example, the change in the case that the opening degree SE of the EGR valve 31 is corrected to set the actual speed to the target speed, and next the opening degree ST of the throttle valve 20 is corrected to adjust the actual negative pressure of the intake pipe to the target negative pressure of the intake pipe. By controlling the speed and the vacuum of the intake manifold, the parameters so early adjusted to the target values.

After the amount of injected fuel has been corrected, as mentioned above, the fuel injection timing is corrected to the pressure change within the combustion chamber 5 to set the target pressure change. Even if the fuel injection timing in this case is corrected only at this stage, it has no great influence on the rotational speed, the suction pipe negative pressure, and the fuel injection amount, so that the necessity of re-controlling to set the parameters to the target values is very poor at best is low.

If The operation of the machine components is thus controlled so that setting a specific parameter to the target value is the Necessity, the already corrected operation of the machine component to correct again, as described above at most extremely low, so that only a very low probability exists that all target values deviate and it comes in the control to chase after. It is therefore possible during the Idle mode the parameter that the controller uses to limit the Machine vibrations needed, early to set the target value.

To perform the control, must be in the in 1 shown expansion tank 12 a pressure sensor for detecting the negative pressure of the intake pipe and within the combustion chamber 5 a combustion pressure sensor for detecting the pressure within the combustion chamber 5 are located. Furthermore, it is preferable between the in 7 shown second operating range II and the first operating range I set a lower limit Z (N) than the limit Y (N) and correct each of the parameters at the time of idling, if the detected value is smaller than the limit Z (N).

13 FIG. 14 shows the target air-fuel ratio when the second combustion mode is performed according to the conventional combustion method. In this case, the curves indicated by A / F = 24, A / F = 35, A / F = 45 and A / F = 60 indicate 13 each states having target air-to-fuel ratios of 24, 35, 45, and 60, respectively. The target opening degree ST of the throttle valve 20 which is needed to set the air-fuel ratio of the target air-fuel ratio is previously in the ROM 42 in the form of a like 14A map shown as a function of the required load L and the rotational speed N, while the target opening degree SE of the EGR control valve 31 which is needed to set the air-fuel ratio to the target air-fuel ratio, previously in the ROM 42 in the form of a like in 14B map shown as a function of the required load L and the rotational speed N is stored.

Further, when the second combustion mode is performed, the injection amount Q is controlled according to the required load L and the rotational speed N, with the injection quantity Q previously in the ROM 42 in the form of a like in 15A map shown as a function of the required load L and the rotational speed N is stored. Moreover, by trial, an optimal injection start timing θS is precalculated when the second combustion mode is performed, and the optimum injection start timing θS is previously in the ROM 42 in the form of a like in 15B map shown as a function of the required load L and the rotational speed N stored.

In the in 12B Thus, the map shown is the optimal injection start timing θS stored according to the first combustion mode. However, as mentioned at the beginning, the optimum injection timing θS changes according to the environmental conditions such as the climate, so that even if the injection start timing is controlled to the optimum injection start timing θS, the fuel injection timing previously determined according to the operating state of the engine and as in 12B has been shown to produce smoke or misfires as environmental conditions change.

If So the injection start time compared to the optimal Injection start timing θS delayed gradually worsens the combustion and eventually misfires are generated. With the himself worsening combustion increases in this case the Torque change of Machine power, making it possible is based on the torque change amount of the engine power to decide whether the combustion has deteriorated or Not. About that In addition, the amount of torque change becomes the engine power with advancing the injection start timing gradually smaller, but soot is generated at the same time when the injection start time is too much preferred. When the torque change amount of engine power controlled within a certain range can be thus achieve a better combustion, in which neither misfires still smoke are generated.

The first embodiment of the invention is now constructed such that an average value of the torque change amount of the engine power is calculated to advance the injection start timing and to prevent the generation of misfires when the average torque change amount is larger than an upper limit value MAX and to retard the injection start timing and to prevent the generation of smoke, though the average torque change amount becomes smaller than a lower limit MIN.

As mentioned above has become, the injection start timing θS on the basis of the torque change amount the engine power controlled. The following is summarized first an example for A method of calculating the torque change amount is described.

For example, if the angular velocity of a crankshaft at the time when the crankshaft is moving from an upper compression dead center (hereinafter referred to as "TDC") to an angle of 30 degrees after the upper compression dead center (hereinafter referred to as "TDC") is designated as a second angular velocity πb, the angular velocity of the crankshaft increases at a time when the crankshaft moves from 60 degrees TDC to 90 degrees TDC, when in each cylinder the angular velocity of the crankshaft increases Combustion occurs because of the combustion pressure from the first angular velocity πa to the second angular velocity πb. When the moment of inertia of the machine motion is set to I, the kinetic energy increases from (1/2) Iπa ² to (1/2) Iπb ² due to the combustion pressure. In summary, since the torque is generated by a higher amount (1/2) I (πb ² -πa ² ) of the kinetic energy, the generated torque is proportional to (πb ² -πa ² ). Accordingly, the generated torque can be calculated from the residual amount between the square of the first angular velocity πa and the square of the second angular velocity πb.

Next, referring to 16 describe a method for calculating the torque generated by each cylinder, illustrating the case that the in 1 shown internal combustion engine is a 4-cylinder internal combustion engine. As mentioned above, the crank angle sensor generates 52 at any time when the crankshaft is around the crank angle 30 Grad rotates, a dispensing pulse and is the crank angle sensor 52 is arranged in such a manner as to generate the output pulse to the top compression dead center OT of a first cylinder C1, a second cylinder C2, a third cylinder C3, and a fourth cylinder C4. Accordingly, the crank angle sensor generates 52 the output pulse for each crank angle of 30 degrees based on the TDC of the cylinders C1, C2, C3 and C4. The firing order in the internal combustion engine used in the invention is 1-3-4-2 in this case.

In 16 the ordinate axis indicates the running time of the crank angle T30 of 30 degrees after the crank angle sensor 52 has generated the output pulse and before it generates the next output pulse. Furthermore, Ta (i) specifies the time of flight from the TDC of the cylinder numbered i to 30 degrees after TDC and Tb (i), the running time of 60 degrees TDC of the cylinder numbered i to 90 degrees after TDC. Thus, for example, Ta (1) indicates the time of flight from the TDC of the first cylinder to 30 degrees after TDC and Tb (1) the running time of 60 degrees TDC of the first cylinder to 90 degrees TDC. When the crank angle of 90 degrees is divided by the transit time T30, the division result also indicates the angular velocity π. Accordingly, crank angle / Ta (i) of 30 degrees indicates the first angular velocity πa in the first cylinder and crank angle / Tb (i) of 30 degrees the second angular velocity πb in the first cylinder.

17 FIG. 15 shows a routine for calculating the amount of torque change in the first embodiment, which is shown in FIG 1 shown and according to the 2 - 15B operated internal combustion engine, this routine is executed per crank angle of 30 degrees by interrupt.

As in 17 is shown first in step 100 decided whether the current position is at 30 degrees to TDC of the cylinder i. If the current position is not 30 degrees TDC of cylinder number i, the process goes to step 102 , where it is decided if the current position is at 90 degrees TDC of cylinder number i. If the current position is not 90 degrees after TDC of cylinder number i, the process cycle is ended.

If in contrast in step 100 it is decided that the current position is at 30 degrees TDC of the cylinder numbered i, the process goes to step 101 and based on the remaining amount between a present time TIME and a time TIMEO before the crank angle of 30 degrees, the running time Ta (i) is calculated from the TDC of the cylinder numbered i to 30 degrees after TDC. If next in step 102 it is decided that the current position is at 90 degrees TDC of cylinder number i, the process goes to step 103 where, based on the balance between the current time TIME and a time TIMEO before the crank angle of 30 degrees, the running time Tb (i) of 60 degrees TDC of the cylinder numbered i to 90 degrees TDC is calculated.

Next will be in step 104 calculated on the basis of the following formula, the generated torque DM (i) of the cylinder with the number i. DN (i) = (πb 2 - πa 2 ) = (30 ° / Tb (i)) 2 - (30 ° / Ta (i)) 2

Next will be in step 105 calculated based on the following formula, the amount of torque change DLN (i) during a cycle in the same cylinder. DLN (i) = DN (i) j - DN (i)

there presses DN (i) j the generated torque in the same cylinder one cycle (a crank angle of 720 degrees) before DN (i).

Next will be in step 106 a count C increased by exactly one. Next will be in step 107 whether the count value C has become 4, that is, whether the torque change amount DLN (i) has been calculated for all the cylinders. If the relationship C = 4 is detected, the process goes to step 108 and, based on the following formula, a torque end amount SM is set to an average value of the torque change amounts DLN (i) in all the cylinders. SM = (DLN (1) + DLN (2) + DLN (3) + DLN (4)) / 4

Next will be in step 109 the count C is set to 0.

Next, referring to 18 a routine for executing the operation control according to the first embodiment will be described. In this case, the routine is executed by each interrupting crank angle.

As in 18 is shown in step 200 First, it is decided whether there is a flag I for the operating range of the machine in the first operating range I. If the flag I is set or the operating range of the machine is in the first operating range I, the process goes to step 201 in which it is decided whether the required load L is greater than the first limit X (N). If the relation L ≦ X (N) holds, the process goes to step 203 and the low-temperature combustion is performed.

And that will be in step 203 based on the in 11A shown map of the target opening degree ST of the throttle valve 20 calculates and becomes the opening degree of the throttle valve 20 set to the target opening degree ST. Next will be in step 204 based on the in 11A shown map of the target opening degree 5E of the EGR control valve 31 calculates and becomes the opening degree of the EGR control valve 31 set to the target opening degree SE. Next will be in step 205 based on the in 12A shown map, the injection quantity Q calculated, wherein next in step 206 on the basis of in 12B an optimal injection start timing θS with respect to the required load L and the rotational speed N is calculated.

Next will be in step 207 decided whether the average torque change amount M is greater than the upper limit value MAX. If the relationship SM> MAX is true, the process goes to step 208 and a fixed value α is added to a correction value ΔθS for the injection start timing θS. Next, the process goes to step 211 and a final injection start timing θS is calculated by adding the correction value ΔθS to the injection start timing θS, which is determined from the in 12B calculated map was calculated. The injection start timing θS is expressed as an advance angle amount, therefore, the condition SM> MAX can be understood to advance the injection start timing θS.

If, in contrast, in step 207 If it is decided that the relation SM ≦ MAX holds, the process goes to step 209 and it is judged whether the average torque change amount SM is smaller than the lower limit MIN. If the relationship SM <MIN holds, the process goes to step 210 and the fixed value α is subtracted from the correction value ΔθS. Next, the process goes to step 211 , Accordingly, at this time, the injection start timing θS is delayed.

If in contrast in step 209 is decided that the relation SM ≥ MIN holds, the process goes to step 211 , Accordingly, when the relation MAX ≥ SM ≥ MIN holds, no renewing operation is performed for the correction value ΔθS.

If, in contrast, in step 201 it is decided that the relationship L> X (N) holds, the process goes to step 202 and the flag I is reset, the process next to step 214 goes and the second combustion mode is performed.

And that will be in step 214 based on the in 14A shown map of the target opening degree ST of the throttle valve 20 calculates and becomes the opening degree of the throttle valve 20 set to the target opening degree ST. Next will be in step 215 based on the in 14B shown map of the target opening degree SE of the EGR control valve 31 calculates and becomes the opening degree of the EGR control valve 31 set to the target opening degree SE. Next will be in step 216 the injection amount Q based on the in 15A calculated map shown next in step 217 on the basis of in 15B shown map of the injection start timing θS is calculated.

If the flag I is reset, the Process in the next process cycle of step 200 to step 212 and it is decided whether the required load L is lower than the second limit Y (N). If the relation L ≥ Y (N) holds, the process goes to step 214 and the second combustion mode is performed at a lean air-fuel ratio.

If, in contrast, in step 212 it is decided that the relation L <Y (N) holds, the process goes to step 213 and the flag I is set, with the process next to step 203 goes and the low-temperature combustion is performed.

When next is a second embodiment of the invention described. The second embodiment of the invention is structured that the amount of torque change the engine power for each of the cylinders is calculated to be the injection start timing of the corresponding cylinder and the emergence of misfires to prevent when the amount of torque change in the respective Cylinder bigger than the upper limit is MAX, and the injection start timing of the corresponding cylinder to delay and the emergence of To prevent smoke when the amount of torque change in each Cylinder is smaller than the lower limit MIN.

19 FIG. 15 shows a routine for calculating the torque change amount in the second embodiment, wherein the routine is executed by cranking at 30 degrees crank angle.

In step 300 from 19 First, it is decided if the current position is at 30 degrees TDC of the cylinder of number i. If the current position is not 30 degrees after TDC of the cylinder of number i, the process jumps to step 302 and it is decided if the current position is at 90 degrees TDC of the cylinder of number i. If the current position is not 90 degrees TDC of cylinder number i, the process cycle is complete.

If in contrast in step 300 it is decided that the current position is at 30 degrees TDC of the cylinder of number i, the process goes to step 301 and, based on a remaining amount between a present time TIME and a time TIMEO before the crank angle of 30 degrees, the running time Ta (i) from the cylinder of the number i is calculated to be 30 degrees after TDC. If next in step 302 it is decided that the current position is at 90 degrees TDC of cylinder number i, the process goes to step 303 and, based on the remaining amount between the present time TIME and a time TIMEO before the crank angle of 30 degrees, the running time Tb (i) is calculated from 60 degrees after TDC of the cylinder of the number i to 90 degrees after TDC.

Next will be in step 304 calculated on the basis of the following formula, the generated torque DN (i) of the cylinder of the number i. DN (i) = (πb 2 - πa 2 ) = (30 ° / Tb (i)) 2 - (30 ° / Ta (i)) 2

Next will be in step 305 calculated based on the following formula, the amount of torque change DLN (i) during a cycle in the same cylinder. DLN (i) = DN (i) j - DN (i)

Here, DN (i) _{j expresses} the generated torque in the same cylinder one cycle (crank angle of 720 degrees) before DN (i).

Next will be in step 306 based on the following formula, a mean value SM (i) of the torque change amount in the cylinder of the number i is calculated. SM (i) = (DLN (i) + DLN (i) j + DLN (i) j-1 + DLN (i) j-2 ) / 4

Here, DLN (i) _j gives the torque change amount of the cylinder of the number i one cycle before, DLN (i) _j-1 the torque change amount of the cylinder number i two cycles before, and DLN (i) _j-2 the torque change amount of the cylinder of the number i three Cycles before.

Next, referring to 20 A routine for executing the operation control according to the second embodiment will be described. In this case, the routine is executed by each interrupting crank angle.

As in 20 is shown in step 400 First, it is decided whether a flag I indicates that the operating range of the machine is in the first operating range I. If the flag I is set or the operating range of the machine is in the first operating range I, the process goes to step 401 in which it is decided whether the required load L is greater than the first limit X (N). If the relation L ≦ X (N) holds, the program goes to step 403 and the low-temperature combustion is performed.

And that will be in step 403 based on the in 11A shown map of the target opening degree ST of the throttle valve 20 calculates and becomes the opening degree of the throttle valve 20 set to the target opening degree ST. Next will be in step 404 based on the in 11B shown map of the target opening degree SE of the EGR control valve 31 calculates and becomes the opening degree of the EGR control valve 31 set to the target opening degree SE. Next will be in step 405 on the basis of in 12A is calculated, the injection amount Q being calculated next in step 406 on the basis of in 12B an optimal injection start timing θS (i) with respect to the required load L and the rotational speed N is calculated.

Next will be in step 407 It is decided whether the average torque change amount SM (i) of the cylinder of the number i is larger than the upper limit MAX. If the relationship SM (i)> MAX holds, the process goes to step 408 and a fixed value α is added to a correction value ΔθS (i) for the injection start timing θS (i) of the cylinder of the number i. Next, the process goes to step 411 and a final injection start timing θS (i) is calculated by adding the correction value ΔθS (i) to the injection start timing θS (i) of the cylinder number i, which is determined from the in 12B calculated map was calculated. As mentioned above, the injection start timing θS is expressed by an advance angle amount, therefore, the condition SM (i)> MAX can be understood to advance the injection start timing θS (i) of the cylinder of the number i.

If, in contrast, in step 407 If it is decided that the relation SM (i) ≦ MAX holds, the process goes to step 409 and it is decided whether or not the average torque change amount SM (i) of the cylinder of the number i is smaller than the lower limit MIN. If the relationship SM (i) <MIN holds, the process goes to step 410 and the fixed value α is subtracted from the correction value ΔθS (i) for the cylinder of the number i. Next, the process goes to step 411 , Accordingly, at this time, the injection start timing θS (i) of the cylinder of the number i is delayed.

If, in contrast, in step 409 is decided that the relation SM (i) ≥ MIN holds, the process goes to step 411 , Accordingly, when the relation MAX ≥ SM (i) ≥ MIN holds, the renewal process of the correction value ΔθS (i) for the cylinder of the number i is not performed.

If, in contrast, in step 401 it is decided that the relationship L> X (N) holds, the process goes to step 402 and the flag I is reset, the process next to step 414 goes and the second combustion mode is performed.

And that will be in step 414 based on the in 14A shown map of the target opening degree ST of the throttle valve 20 calculates and becomes the opening degree of the throttle valve 20 set to the target opening degree ST. Next will be in step 415 based on the in 14B shown map of the target opening degree SE of the EGR control valve 31 calculates and becomes the opening degree of the EGR control valve 31 set to the target opening degree SE. Next will be in step 416 on the basis of in 15A shown map, the injection quantity Q calculated, wherein next in step 417 based on the in 15B shown map of the injection start timing θS (i) is calculated.

If flag I is reset, the process goes off in the next process cycle 400 to step 412 and it is decided whether the required load L is lower than the second limit Y (N). If the relation L ≥ Y (N) holds, the process goes to step 414 and the second combustion mode is performed under a lean air-fuel ratio.

If, in contrast, in step 412 it is decided that the relation L <Y (N) holds, the process goes to step 413 and the flag I is set, with the process next to step 403 goes and the low-temperature combustion is performed.

apart of which, according to the first / second embodiment the invention delayed by the fixed value α or when the detected amount of torque change is below the lower limit MIN or beyond the upper limit MAX, however, the injection timing is not limited to this control. Of the Injection timing can also be controlled in any other way if the detected amount of torque change is taken into account place. For example, control of the injection timing may be dependent on from the detected torque change amount such that the injection timing is advanced or retarded, if the detected amount of torque change is not within of the predetermined area.

As in the above mentioned 9 is shown, in the case of switching from the first combustion mode to the second combustion mode, it is necessary to determine the injection amount in addition to the above-mentioned control of the injection timing, so that I do not change the output torque of the engine when the injection amount is suddenly lowered.

It will now be the content of the supplementary Injection amount control at the time of switching the combustion state described.

The second combustion mode has a slightly better thermal efficiency than the first combustion mode, and therefore the fuel injection amount required to achieve the same engine output torque is lower in the second combustion mode than in the first combustion mode. Since the injection amount at the same depression amounts of the accelerator pedal 50 and the same speed N is such that the in 21A shown injection amount Q1 when performing the first combustion mode in the entire operating range except for the low-speed range greater than that in 21B is shown injecting amount Q2 when performing the second combustion mode, the injection amount Q1 becomes when performing the first combustion mode in the ROM 42 in the form of a like in 22A shown map as a function of the speed N and the depression amount of the accelerator pedal 50 (or the required load) is stored and the injection amount Q2 when performing the second combustion mode in the ROM 42 in the form of a like in 22B shown map as a function of the speed N and the depression amount of the accelerator pedal 50 (or the required load) stored. The injection quantities Q1 and Q2 may be in ROM 42 also be stored in the form of the function.

The solid lines in the 21A and 21B In this case, each indicate a line connecting the points at which the amount of depression of the accelerator pedal respectively 50 is equal to and presses the numbers on each solid line the amount of depression of the accelerator pedal 50 in percent.

As further in 21B is shown, the injection amount Q2 is suddenly increased in the second combustion mode with respect to a same amount of depression of the accelerator pedal with decreasing speed N, to prevent stalling of the engine with decreasing speed N. However, if the injection amount is suddenly increased in the first combustion mode with decreasing speed N, the air-fuel ratio becomes excessively dense or rich. If the air-fuel ratio becomes excessively rich, misfire will occur due to lack of air, making it easier to stall the engine. Now if the first combustion mode as in 21A is performed, the structure is such that the rising speed of the injection amount Q1 in the low rotation region with the engine speed N lower than that in FIG 21B shown second combustion mode, so that the air-fuel ratio is not excessively rich.

When performing the first combustion mode, the injection amount Q1 is therefore always based on the in 22A In the performance of the second combustion mode, the injection amount Q2 is calculated on the basis of the map shown in FIG 22B is calculated, so that the injection amount Q is immediately lowered when the operating state of the engine as in 23A shown moved from the first operating range I to the second operating range II, and the injection amount Q is increased immediately when the operating state of the machine as in 23B shown moved from the second operating region II to the first operating region I. Moreover, it is also possible to increase and decrease the injection amount from the switching timing of the operating state after a predetermined time delay. Thereby, it is possible to reduce the change of the output torque at the time of switching the combustion state.

The above-mentioned embodiment is constructed such that the fuel injection amount is calculated on the basis of the depression amount of the accelerator pedal and the rotational speed. Specifically, the structure is such that the fuel injection amount is previously stored in the form of a map as a function of the depression amount of the accelerator pedal and the rotational speed, and the fuel injection amount is calculated from the map. However, in the calculation of the fuel injection amount with the aid of the above-mentioned map, there is the risk that the injection amount is difficult to adjust when the fuel injection amount at the time of switching from the first combustion mode to the second combustion mode has to be reduced. The reason for this is now referring to 24 described. In 24 the reference character L designates the depression amount of the accelerator pedal, the reference character N the rotational speed, and the reference character X the boundary between the operation region I in which the first combustion mode is performed and the operation region II in which the second combustion mode is performed. As in 24 is shown, it is difficult in this case to set the fuel injection amount close to the limit X, when the structure is such that the fuel injection amount Q1 for the operating region I in the form of a map as a function of the depression amount L of the accelerator pedal and the rotational speed N stored and also the fuel injection quantity Q2 for the operating region II is stored in the form of a characteristic map as a function of the depression amount L of the accelerator pedal and the rotational speed N.

Namely, it is important that the generation of a torque difference be prevented when the required load L exceeds the limit X, which is necessary for achieving this effect, the fuel injection amount Q1 and the fuel injection amount Q2 at all points on the boundary X in such a manner that the same torque is generated at all points on the boundary X. However, in order to calculate the fuel injection amounts Q1 and Q2 in which the torque is equal for each of the points on the boundary X, the fuel injection quantities Q1 and Q2 for each of the points on the boundary X must be set on the basis of an experiment so that quite difficult to calculate the fuel injection quantities Q1 and Q2 on the boundary X.

Of the Structure is therefore such that the required torque of the machine is calculated and the fuel injection amount based calculated calculated required torque of the speed when either the first combustion mode or the second Combustion mode performed becomes. Accordingly, the output torque can be prevented be at the time of switching between the first combustion mode and the second combustion mode without being complicated Work would be necessary.

With reference to 25 Firstly, a first operating region I, in which the first combustion mode or the low-temperature combustion is performed, and a second operating region II, in which the second combustion mode or the combustion according to the conventional method is performed. The ordinate axis TQ in 25 In this case, the required torque and the abscissa axis N indicate the speed. In addition, X (N) is in 25 a first boundary between the detected operation area I and the second operation area II, while Y (N) indicates a second boundary between the first operation area I and the second operation area II. On the basis of the first boundary X (N), a change of the first operating region from the first operating region I to the second operating region II is decided, while on the basis of the second boundary Y (N) a change of the operating region from the second operating region II to the first operating region I is decided becomes.

If the required torque TQ beyond the function the speed N corresponding to the first limit X (N) and if the operating state of the machine corresponds to the first operating range I. and the low-temperature combustion is carried out, it is decided that the operating range moves to the second operating range II, so that the combustion is carried out according to the conventional combustion method. If the required torque TQ is now below that of a function the speed N corresponding to the second boundary Y (N), it is decided that the operating range goes to the first operating range I, so that again the low-temperature combustion is carried out, the two above Limits, the first limit X (N) and the second, closer to lower torques lying limit Y (N), there are the following two reasons. Of the The first reason is that because of that, because the combustion temperature in the second operating range II on the high load side relatively high is, the low-temperature combustion can not be done immediately, even if the required load L at this time under the first boundary X (N) falls. That's because the low temperature combustion only then begins when the required load L significantly lower or lower than the second boundary Y (N). The second reason is, that in terms of change the operating range between the first operating range I and the second operating area II for a hysteresis is taken care of.

26A FIG. 14 shows the relationship between the required torque TQ, the amount of depression L of the accelerator pedal 50 and the speed N. In 26A In this case, each curve expresses a constant torque, where the curve labeled TQ = 0 indicates that the torque 0 and the other curves indicate that the required torque increases gradually in the order TQ = a, TQ = b, TQ = c and TQ = d. This in 26A required torque TQ shown previously in ROM 42 in the form of a like in 26B shown map as a function of the depression amount L of the accelerator pedal 50 and the speed N stored. The required torque TQ, the amount of depression L of the accelerator pedal 50 and the rotational speed N, according to the invention, first based on the in 26B is calculated, and then based on the required torque TQ, the fuel injection amount Q and the like are calculated.

27 shows the air-fuel ratio A / F in the first operating region I. In 27 the curves labeled A / F = 15.5, A / F = 16, A / F = 17 and A / F = 18 indicate states with air-fuel ratios of 15.5, 16, 17 and 18 respectively, wherein the air-fuel ratios between the curves are each defined by proportional allocation. As in 27 is shown, the air-fuel ratio in the first operating region I is lean and the air-fuel ratio A / F is also made lean in the first operating region I in accordance with the decreasing required load L.

With decreasing torque required TQ so the heat of combustion is lowered. Accordingly, the low-temperature combustion can be performed even when the EGR ratio is lowered as the required torque TQ decreases. With decreasing EGR ratio The air-fuel ratio increases, so that the air-fuel ratio A / F as in 27 shown increases with decreasing required torque TQ. With increasing air-fuel ratio A / F, the specific fuel consumption improves so that the air-fuel ratio A / F according to the embodiment of the present invention is increased with decreasing required torque TQ to make the air-fuel ratio as lean as to make possible.

28A shows the injection amount Q in the first operating range I and 28B the injection start timing .theta.S in the first operating region I. As in FIG 28A is shown, the injection quantity Q in the first operating region I before in ROM 42 is stored in the form of a map as a function of the required torque TQ and the speed N and is, as in 28B is shown, the injection start point θs in the first operating region I previously in the ROM 42 stored in the form of a map as a function of the required torque TQ and the rotational speed N.

In addition, the target opening degree ST of the throttle valve 20 needed to set the air-fuel ratio to the in 27 set target air-fuel ratio previously set as in 29A shown in the ROM 42 stored in the form of a map as a function of the required torque TQ and the rotational speed N, while the target opening degree SE of the EGR control valve 31 needed to set the air-fuel ratio to the in 27 set target air-fuel ratio previously set as in 29B shown in the ROM 42 in the form of a map as a function of the required torque TQ and the rotational speed N is stored.

Further, in the embodiment of the present invention, the fuel injection pressure in the first operating region I and the target fuel pressure P within the common rail, respectively 34 previously as in 29C shown in ROM 42 in the form of a map as a function of the required torque TQ and the speed N stored.

30 FIG. 14 shows the target air-fuel ratio when the second combustion mode is performed according to the conventional combustion method.

In 14 in this case indicate the curve states with target air-fuel ratios of 24, 35, 45 and 60, designated A / F = 24, A / F = 35, A / F = 45 and A / F = 60.

31A shows the injection quantity Q in the second operating range II and 31B the injection start timing θS in the second operating region II. As in FIG 31A is shown, the injection amount Q in the second operating region II before in ROM 42 is stored in the form of a map as a function of the required torque TQ and the rotational speed N, and the injection start timing θS in the second operating region II before as in 31B shown in the ROM 42 stored in the form of a map as a function of the required torque TQ and the rotational speed N.

Further, the target opening degree ST of the throttle valve becomes 20 needed to set the air-fuel ratio to the in 30 set target air-fuel ratio previously set as in 32A shown in the ROM 42 stored in the form of a map as a function of the required torque TQ and the rotational speed N, while the target opening degree SE of the EGR control valve 31 needed to set the air-fuel ratio to the in 30 set target air-fuel ratio previously set as in 32B shown in the ROM 42 in the form of a map as a function of the required torque TQ and the rotational speed N is stored.

Further, according to this embodiment, a fuel injection pressure in the first operating region II and a target fuel pressure P within the common rail, respectively 34 previously as in 32C shown in the ROM 42 stored in the form of a map as a function of the required torque TQ and the rotational speed N.

According to this embodiment, the injection amount Q in the first operating region is calculated on the basis of in 28A shown map and the injection quantity Q in the second operating range II on the basis of in 31A calculated map shown. The in the 28A and 31A maps shown are each a function of the required torque TQ and the rotational speed N, therefore, the required torque does not change even at the time of switching when the combustion at the first boundary X (N) or the second boundary Y (N) from the first Combustion mode is switched to the second combustion mode or from the second combustion mode to the first combustion mode. Accordingly, there is no risk that a torque change is generated at the time of switching.

In the 28A shown injection quantity Q for the first combustion mode and in 31A shown injection amount Q for the second combustion mode are calculated in this case by a trial. Since the injection quantity Q for the first combustion mode and the injection quantity Q for the second combustion mode are each independently calculated by a trial without the torque differences If we take into account the two limits X (N) and Y (N), we can add the ones in the 28A and 31A form maps shown in this case in a much easier way.

In contrast, this is indicated by the curves in 26A specified torque each experimentally set so that the vehicle better driving behavior can be ensured. As a result, it is the case that the actual torque output by the engine is on the in 26A with TQ = 0 is not 0. Therefore, according to this embodiment, the construction is such that it is decided whether an operating condition in which the engine load is largely O, is a value obtained by subtracting a shift amount of the required torque TQ with respect to 0 from that of in 26A calculated torque required torque TQ is achieved, is set to an actual required torque, if based on the in 26A at a time when it is judged that the operating state in which the engine load is largely 0, is shifted to the positive side with respect to 0, and that a value obtained by adding the shift amount of the required torque TQ with respect to 0 to that on the basis of in 26A calculated torque required torque TQ is achieved, is set to the actually required torque, if based on the in 26A At the time when it is judged that the operating condition in which the engine load is largely zero is calculated, the required torque TQ calculated is shifted to the negative side with respect to zero.

For example, when the operating state at a torque of 0 actually output from the engine is 0, for example 26A at the point F shown in Fig. 10, in which the low-pressure amount of the accelerator pedal L ₀ and the rotational speed N ₀ , the required torque TQ at that time in the in 26B is not 0 and the required torque TQ in the map is shifted from 0 by an amount ΔKG to the positive side. In this case, in the embodiment of the present invention, the value (TQ-ΔKG) obtained by subtracting the shift amount ΔKG of the required torque from that based on the in 26B calculated required torque TQ is set to the actual required torque. As a result, it is possible with respect to each low-pressure amount L of the accelerator pedal 50 and each speed N to achieve the originally intended engine output torque.

Next, referring to 33 described the operation control.

In step 500 from 33 It is first decided whether the flag I indicating the operating range of the machine is in the first operating range I. If the flag I is set or the operating range of the machine is in the first operating range I, the process goes to step 501 to decide whether the required load L is greater than the first limit X (N). If the relation L ≦ X (N) holds, the process goes to step 503 and the low-temperature combustion is performed.

And that will be in step 503 the required torque TQ on the basis of in 26B calculated map shown. Next will be in step 504 the value (= TQ - ΔKG) obtained by subtracting the required torque shift amount ΔKG from the required torque TQ calculated based on the map is set to a final required torque TQ. Next, in the steps 505 to 509 using the final required torque, the opening degree ST of the throttle valve, the opening degree SE of the RGR control valve, the injection amount Q, the injection start timing θS, and the fuel injection pressure P are calculated.

And that will be in step 505 on the basis of in 29A shown map of the target opening degree ST of the throttle valve 20 calculates and becomes the opening degree of the throttle valve 20 set to the target opening degree ST. Next will be in step 506 on the basis of in 29B shown map of the target opening degree SE of the EGR control valve 31 calculates and becomes the opening wheel of the EGR control valve 31 set to the target opening degree SE. Next will be in step 507 on the basis of in 28A shown map, the injection quantity Q calculated. Next will be in step 508 the injection start timing θS based on the in 28B calculated map shown. Next will be in step 509 on the basis of in 29C The map shows the target fuel pressure within the common rail 34 or the injection pressure P calculated.

Next will be in step 510 based on the output of the idle switch 55 decided whether the throttle valve 20 to open the idle opening degree. When the throttle valve 20 to open the idle opening degree, the process goes to step 511 and is based on the output pulse of the vehicle velocity sensor 350 decided whether the vehicle speed is 0. If the vehicle speed is 0, the process goes to step 512 and is based on the output of the neutral switch 54 decided whether the automatic transmission 38 in the neutral position. When the automatic transmission 38 in the neutral position, the process goes to step 513 and it is decided if a fixed period of time has elapsed after the automatic transmission 38 has come to the neutral position, and it is decided after the lapse of the fixed period that the operating state is reached, in which the engine load is largely zero.

If it is decided that the operating condition is reached in which the engine load is generally 0, the process goes to step 514 in which the required torque TQ on the basis of in 26B is calculated, and is next in step 515 the required torque TQ is set to the shift amount ΔKG.

If, in contrast, in step 501 it is decided that the relationship L> X (N) holds, the process goes to step 502 and the flag I is reset, the process next to step 518 goes and the second combustion mode is performed.

And that will be in step 518 the required torque TQ on the basis of in 26B calculated map shown. Next will be in step 519 the value (= TQ - ΔKG) obtained by subtracting the required torque displacement contract ΔKG from the required torque TQ calculated based on the map is set to the final required torque TQ. Next, in the steps 520 to 524 using the final required torque, the opening degree ST of the throttle valve, the opening degree SE of the EGR control valve, the injection amount Q, the injection start timing θS, and the fuel injection pressure P are calculated.

And that will be in step 520 on the basis of in

32A shown map of the target opening degree SE of the throttle valve 20 calculates and becomes the opening degree of the throttle valve 20 set to the target opening degree ST. Next will be in step 521 on the basis of in 32B shown map of the target opening degree SE of the EGR control valve 31 calculates and becomes the opening degree of the EGR control valve 31 set to the target opening degree SE. Next will be in step 522 on the basis of in 31A shown map, the injection quantity Q calculated. Next will be in step 523 on the basis of in 32B shown map, the injection start timing θS calculated. Next will be in step 524 the target fuel pressure within the common rail 34 or the injection pressure P on the basis of in 32C calculated map shown.

If the flag I is reset, the process goes to the steps in the next process cycle 500 to 516 and it is decided whether the required load L is smaller than the second limit Y (N). If the relation L> Y (N) holds, the process goes to step 518 and the second combustion mode is performed under the lean air-fuel ratio. If, in contrast, in step 516 it is decided that the relation L <Y (N) holds, the process goes to step 517 and the flag I is set, with the process next to step 503 goes and the low-temperature combustion is performed.

The The above description of the preferred embodiments is not restriction of the claimed invention and the discussed combination of features must for the invention solution not necessarily necessary.

A first combustion mode or a low-temperature combustion, in which the amount of an EGR gas within a combustion chamber ( 5 ) is larger than the amount of the EGR gas when a generation amount of soot peaks and hardly generates soot, and a second combustion mode is performed in which the amount of EGR gas within the combustion chamber (FIG. 5 ) is smaller than the amount of EGR gas when the production amount of the soot becomes the highest value. It is characterized by a detection device ( 52 . 40 ) detects a torque change amount of engine power when the first combustion mode is performed, wherein the fuel injection timing is controlled so that the torque change amount falls within a predetermined range.

A first combustion mode or a low-temperature combustion, in which the amount of an EGR gas within a combustion chamber ( 5 ) is larger than the amount of the EGR gas when a generation amount of soot peaks and hardly generates soot, and a second combustion mode is performed in which the amount of EGR gas within the combustion chamber (FIG. 5 ) is smaller than the amount of EGR gas when the production amount of the soot becomes the highest value. It is characterized by a detection device ( 52 . 40 ) A torque change amount of the engine power he when the first combustion mode is performed, wherein the fuel injection timing is controlled so that the torque change amount falls within a predetermined range.

Claims (10)

An internal combustion engine constructed such that a generation amount of soot gradually increases to a maximum value as the amount of an inert gas supplied within a combustion chamber increases, comprising: an inert gas amount controller (10); 40 ) for restricting the temperature of a fuel and an ambient gas at the time of combustion within the combustion chamber to a temperature lower than a temperature when soot is generated by adjusting the amount of inert gas supplied within the combustion chamber to more than the amount of the inert gas when the generation amount of the soot becomes the maximum value is characterized by detecting means (Fig. 52 . 40 ) for detecting a torque change amount of the engine power; and an injection timing control device ( 40 ) for controlling a fuel injection timing so that the torque change amount is within a predetermined range. Internal combustion engine according to Claim 1, in which the injection timing control device ( 40 ) controls the fuel injection timing in consideration of the detected torque change amount. Internal combustion engine according to Claim 1 or 2, in which the injection timing control device ( 40 ) controls the fuel injection timing in accordance with the detected torque change amount so that the injection timing is advanced or retarded when the detected torque change amount is not within the predetermined range. Internal combustion engine according to one of claims 1 to 3, in which the injection timing control device ( 40 ) prefers the fuel injection timing when the torque change amount becomes larger than the predetermined range and delays the fuel injection timing when the torque change amount becomes smaller than the predetermined range. Internal combustion engine according to one of claims 1 to 4, in which the detection device ( 52 . 40 ) detects a mean amount of torque change of all cylinders and the injection timing control means ( 40 ) controls the fuel injection timing of each cylinder so that the mean torque change amount is within the predetermined range. Internal combustion engine according to one of claims 1 to 4, in which the detection device ( 52 . 40 ) detects the amount of torque change on each cylinder and the injection timing control means (FIG. 40 ) controls the fuel injection timing at each cylinder so that the torque change amount of each cylinder is within the predetermined range. Internal combustion engine according to one of Claims 1 to 6, in which a catalytic converter (within the exhaust tract of the engine) ( 25 ) is arranged with oxidation function, which is preferably formed by an oxidation catalyst, a three-way catalyst and / or a NOx absorber. Internal combustion engine according to one of claims 1 to 7, in which an exhaust gas recirculation device ( 29 . 30 . 31 ) is provided to one of the combustion chamber ( 5 ) discharged exhaust gas with preferably an exhaust gas recirculation rate of more than 55% in an engine intake tract ( 17 ), and the inert gas is formed by the recirculated exhaust gas. Internal combustion engine according to one of claims 1 to 8, in which a switching device ( 40 ) is provided for selectively switching between a first combustion mode and a second combustion mode, wherein in the first combustion mode, the amount of an inert gas within the combustion chamber ( 5 ) is larger than the amount of an inert gas when a generation amount of soot peaks and hardly generates soot, and wherein in the second combustion mode, the amount of inert gas within the combustion chamber ( 5 ) is smaller than the amount of the inert gas when the production amount of soot peaks. Internal combustion engine according to claim 9, wherein the Operating range of the machine in a first operating range I close one side of low load and a second operating region II close a high load side, the first combustion mode in the first operating region I and the second combustion mode in the second operating area II.